Method for controlling internal combustion engine valve operating mechanism

ABSTRACT

The reciprocating valve actuation and control system includes a poppet valve moveable between a first and second position; a source of pressurized hydraulic fluid; a hydraulic actuator including an actuator piston coupled to the poppet valve and reciprocating between a first and second position responsive to flow of the pressurized hydraulic fluid to the hydraulic actuator; an electrically operated valve controlling flow of the pressurized hydraulic fluid to the actuator; and an engine computer that generates electrical pulses to control the electrically operated valve. The electrically operated valve preferably comprises a three path rotary latched magnetic motor actuating a rotary valve portion having a housing, a rotor, and a stator receiving and supplying hydraulic fluid pressure to the rotor, which alternately directs the hydraulic fluid pressure to the valve cylinder for opening of the valve, or to return to the engine oil sump, for closing the valve. In a presently preferred embodiment, the hydraulic actuator comprises a self-contained cartridge assembly including an actuator piston with dampers for damping motion of the actuator piston, limiting the actuator stroke to assure soft seating of the actuator, and to avoid overshoot during the engine valve opening stroke and the engine valve return stroke. The electro-hydraulic valves are electrically controlled by the engine computer, which generates electrical signals carried to the electro-hydraulic valves. The engine computer typically senses conventional engine variables, and optimizes performance of the valve actuation and control system according to preestablished guidelines, with information being supplied to the engine computer by sensors. The engine computer controls all aspects of engine performance, interfaces with all of the peripheral sensors, and calculates fuel parameters, ignition timing and engine valve timing based upon prior mapping of the engine. In this manner the engine can be controlled so as to provide maximum fuel economy, minimum emissions, maximum engine torque, or a compromise between these parameters.

BACKGROUND OF THE INVENTION

[0001] 1. Field of the Invention

[0002] This invention relates generally to a valve actuating apparatusfor engines, and more particularly concerns a system for actuating andcontrolling reciprocating valves for the cylinders of an internalcombustion engine.

[0003] 2. Description of Related Art

[0004] Conventional piston type internal combustion engines typicallyutilize mechanically driven camshafts for operation of intake andexhaust valves, with fixed valve lift and return timing and duration.Electrically or hydraulically controlled valves for improved control ofvalve operation have also been used in order to improve fuel economy andreduce exhaust emissions.

[0005] For example, a variable engine valve control system is known inwhich each of the reciprocating intake or exhaust valves ishydraulically controlled, and includes a piston receiving fluid pressureacting on surfaces at both ends of the piston. One end of the piston isconnected to a source of high pressure hydraulic fluid, while the otherend of the piston can be connected to a source of high pressurehydraulic fluid or a source of low pressure hydraulic fluid, under thecontrol of a rotary hydraulic distributor coupled with solenoid valves.

[0006] Another engine valve actuating system is known in which eachcylinder is provided with a coaxial venturi shaped duct having inwardlyfacing vanes that hold an electro-mechanical valve actuator. When theelectro-mechanical valve actuator receives a pulsed electrical signal,the actuator operates to reciprocate the valve.

[0007] While a camshaft driven intake or exhaust valve will typicallyopen and close with a constant period as measured in crankshaft degrees,for any given engine load or rpm, there is a need for an indirect valveactuation system for internal combustion engines that can operate morerapidly, and that will open the valve at the same rate regardless ofengine operating conditions. Ideally, a valve actuation system shouldmatch the optimum, maximum valve rate of operation at maximum speed ofoperation of an engine to provide a rapid, optimum valve operation rate.It would also be desirable to provide a valve actuation system forinternal combustion engines offering a speed of operation that willallow greater flexibility in programming valve events, resulting inimproved low speed torque, lower emissions, and better fuel economy.Conventional approaches to providing higher rates of valve opening andclosing have used non-latching control valves commonly involving systemsusing either spool valves or poppet valves, neither of which provide fora high flow open area in a small, low inertia system or energy efficientlatching mechanisms. It would be desirable to provide a valve actuationand control system with an electro-hydraulic valve system, having a highflow open area, low inertia of operation, a small size, and ease ofmanufacture. The present invention meets these needs.

SUMMARY OF THE INVENTION

[0008] Briefly, and in general terms, the present invention provides foran intake/exhaust (I/E) reciprocating valve actuation and control systemfor the cylinders of an internal combustion engine, comprising I/Epoppet valves moveable between a first and second position; a source ofpressurized hydraulic fluid; a hydraulic actuator including an actuatorpiston coupled to the poppet valve and reciprocating between a first andsecond position responsive to flow of the pressurized hydraulic fluid tothe hydraulic actuator; an electrically operated hydraulic valvecontrolling flow of the pressurized hydraulic fluid to the hydraulicactuator; and electronic control means generating electrical pulses tocontrol the electrically operated valve.

[0009] In one presently preferred embodiment, the invention provides fora three way electrically operated valve controlling flow of thepressurized hydraulic fluid to the actuator, supplying pressure whenelectrically pulsed to open, magnetically latching, and dumping actuatoroil to an engine oil sump when the valve is electrically pulsed toclose. The electrically operated valve preferably comprises a three pathrotary latched magnetic motor actuating a rotary valve portion having ahousing, a rotor, and a stator receiving and supplying hydraulic fluidpressure to the rotor, which alternately directs the hydraulic fluidpressure to the valve cylinder for opening of the valve, or to return tothe engine oil sump, for closing the valve.

[0010] In a presently preferred embodiment, the hydraulic actuatorcomprises a self-contained cartridge assembly including an actuatorpiston with means for damping motion of the actuator piston, limitingthe actuator stroke to assure soft seating of the I/E valve, and toavoid overshoot during the engine valve opening stroke and the enginevalve return stroke. In a currently preferred embodiment, the source ofpressurized hydraulic fluid comprises an engine-driven pump supplyingengine oil under pressure as the hydraulic fluid, an accumulator is usedto provide a reservoir of high pressure fluid, and an engine oil sumpfor receiving return hydraulic fluid. An unloader valve limiting pumpoutput pressure is also provided, along with a check valve preventingbackflow from the engine oil sump. An accumulator is also preferablyprovided for storing a sufficient volume of pressurized hydraulic fluidto moderate the pump and unloader valve duty cycle. The unloader valvepreferably comprises a pressure sensing valve that senses pump outputpressure and opens when the pressure reaches a preset value, so thatwhen the unloader valve is open, flow from the pump returns to theengine oil sump. The accumulator is also used to store energy primarilydissipated under deceleration by the brakes or as a compression brake byfilling the accumulator during that time. The engine would use thetorque from the wheels in reverse driving the hydraulic pump and fillingthe accumulator, thus recycling velocity energy that would normally belost to wheel braking.

[0011] Thus, the hydraulic pump could be temporarily disconnected sothat under high load, the valve train would run off stored accumulatorenergy. This would use more of the power lost during braking. In apresently preferred embodiment, the control means comprises a computer,and sensors are operatively connected to the computer, for monitoringengine variables, and for optimizing performance of the system.

[0012] These and other aspects and advantages of the invention willbecome apparent from the following detailed description and theaccompanying drawings, which illustrate by way of example the featuresof the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

[0013]FIG. 1 is a schematic diagram of the internal combustion enginereciprocating valve actuation and control system of the invention;

[0014]FIG. 2 is a sectional view of a first embodiment of a hydraulicactuator of the reciprocating valve actuation and control system of FIG.1;

[0015]FIG. 3 is a sectional view of a second embodiment of a hydraulicactuator of the reciprocating valve actuation and control system of FIG.1;

[0016]FIG. 4 is a sectional view of a damping spacer of the hydraulicactuator of FIG. 3;

[0017]FIG. 5A is a sectional view of a third embodiment of a hydraulicactuator of the reciprocating valve actuation and control system of FIG.1;

[0018]FIG. 5B is a plan view of the split ring of the hydraulic actuatorof FIG. 5A;

[0019]FIG. 6 is a sectional view of a fourth embodiment of a hydraulicactuator of the reciprocating valve actuation and control system of FIG.1;

[0020]FIG. 7A is a sectional view of a fifth embodiment of a hydraulicactuator of the reciprocating valve actuation and control system of FIG.1;

[0021]FIG. 7B is a plan view of the laminar sealing ring of thehydraulic actuator of FIG. 7A;

[0022]FIG. 7C is a side elevational view of the laminar sealing ring ofFIG. 7B;

[0023]FIG. 8 is a sectional view of the electrically operated valvecontrolling flow of the pressurized hydraulic fluid to the actuator ofthe reciprocating valve actuation and control system of FIG. 1;

[0024]FIG. 9 is a cross-sectional view of the electrically operatedvalve motor taken along line 9-9 of FIG. 8;

[0025]FIG. 10 is a plan view of the rotor of the rotary valve of theelectrically operated valve of FIG. 8;

[0026]FIG. 11 is a sectional view of the rotor taken along line 11-11 ofFIG. 10;

[0027]FIG. 12 is a sectional view of the stator of the rotary valve ofthe electrically operated valve of FIG. 8;

[0028]FIG. 13 is a cross-section of the stator taken along line 13-13 ofFIG. 12;

[0029]FIG. 14 is a sectional view of the rotary valve assembly of theelectrically operated valve of FIG. 8;

[0030]FIG. 15 is a cross-sectional view of the rotary valve assemblytaken along line 15-15 of FIG. 14;

[0031]FIG. 16 is a perspective view of the rotary latched magnetic motorof the electrically operated valve of FIG. 8;

[0032]FIG. 17 is a schematic front view of the rotary latched magneticmotor of FIG. 16, illustrating operation of the motor;

[0033]FIG. 18 is a graph comparing operating speeds of valves driven bya mechanical camshaft and valves driven by the reciprocating valveactuation and control system of the invention; and

[0034]FIG. 19 is a schematic diagram of paired intake and exhaust valvesof unequal sizes.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

[0035] While mechanical camshafts for the intake and exhaust valves ofinternal combustion engines typically have a period of opening andclosing that remains constant in terms of crankshaft degrees for anyengine load or rpm; this has limited the ability of the automotiveindustry to improve fuel economy, reduce harmful exhaust emissions, andto improve low end torque. Typical approaches to providing variablevalve opening and closing positions have involved either variablemechanical linkages or phasing by motors connecting the camshaft to thecam drive. These methods do not provide a high flow open area in a smalllow inertia system.

[0036] The present invention accordingly provides for an improvedreciprocating valve actuation and control system for the cylinders of aninternal combustion engine. As is illustrated in the drawings, and as isgenerally shown in FIG. 1, the reciprocating valve actuation and controlsystem of the invention is a camless valve control system 20 for anengine poppet valve 22 moveable between a first, open position, and asecond, closed position in which the engine poppet valves are reseatedby common valve springs 24. The engine poppet valves are driven byhydraulic actuators 26, which are controlled by electrically operatedelectro-hydraulic valves 28 supplying hydraulic fluid to the actuatorsvia conduit 29. The hydraulic fluid is preferably engine oil, suppliedto the electro-hydraulic valves by the pressure rail 30. An enginedriven hydraulic pump 32 supplies the oil pressure that is used to openthe engine, poppet valves, receiving the oil from an engine oil sump 34.In a presently preferred embodiment, the electro-hydraulic valves arethree way type hydraulic valves, supplying pressure when electricallypulsed to open, magnetically latching, and dumping the actuator oil tothe sump when pulsed to close. Each engine I/E valve is preferablyprovided with an actuator and an electro-hydraulic valve.

[0037] In a presently preferred embodiment, the engine driven pump 32 isa hydraulic pump driven directly by the engine, so that the output ofthe pump will increase in direct proportion to the engine speed. Thepositive displacement pump is preferably sized to provide about 110% ofthe oil flow required by the engine system of valves. The engine oilreturn from the electro-hydraulic valve and piston actuator assembly isto the engine oil sump, typically by gravity through the normal enginedrainage passage (not shown). The positive displacement pump outputpressure is also preferably limited by an unloader valve 36, asmoderated by an accumulator 38 connected to the oil pressure rail. Thenature of the actuator and the valve utilizing the normal engine oilsupply allows the engine oil supply to be used as a hydraulic fluid evenif the engine oil supply contains some entrained air, drasticallysimplifying the system and accessories that would otherwise be requiredto condition the hydraulic fluid, and obviating the need for a separatehydraulic fluid supply.

[0038] The unloader valve 36 preferably comprises a pressure sensingvalve that senses pump output pressure and opens when the pump outputpressure reaches a preset threshold value. When the unloader valve isopened, all of the flow from the positive displacement pump is to returnto the engine oil sump, so that the output from the pump is then“unloaded”. A check valve 40 is also preferably provided in the fluidline between the accumulator and the unloader valve to prevent backflowfrom the accumulator.

[0039] The accumulator in the system is provided to receive oil from thepump, accepting a volume of engine oil from the pump as an accumulatorpiston 42 moves within in the accumulator to create the interioraccumulator volume. A means for biasing the piston to maintain pressureon the piston is also provided, preferably in the form of a coil spring44, although other means of biasing the piston to provide system oilpressure could also be used, such as a pneumatic pressure chamber, forexample. When the unloader valve senses that pump output pressure hasreached the preset threshold value, opening to allow flow from the pumpto return to the engine oil sump, the hydraulic fluid flow and pressureare supplied to the system from the accumulator. When this supply isexhausted, the system pressure drops, the unloader valve senses thesystem pressure drop below a lower, preset minimum oil pressurethreshold, and closes, allowing the pump to reload the accumulatorvolume. The cycling rate of this action depends on the settings of theminimum and maximum oil pressure thresholds of the unloader valve. Theunloader valve settings can be relatively close together, so that thesystem cycles rapidly, or can be set relatively far apart, so that thecycle rate is slower, and resulting in a greater variation of hydraulicfluid supply pressure, as desired. Unloader valve settings can becontrolled by the engine control unit (ECU), or engine computer 50.

[0040] The electro-hydraulic valves are preferably electricallycontrolled by the engine computer 50 (ECU), which generates electricalsignals carried to the electro-hydraulic valves via electricalconnectors 52 a-d. The engine computer typically senses conventionalengine variables, and optimizes performance of the valve actuation andcontrol system according to preestablished guidelines, with informationbeing supplied to the engine computer by sensors 54 a-c. The valveactuation and control system typically includes a manifold pressuresensor, a manifold temperature sensor, a mass flow sensor, a coolanttemperature sensor, a throttle position sensor, an exhaust gas sensor, ahigh resolution engine position encoder, a valve/ignition timing decodercontroller, injection driver electronics, valve coil driver electronics,ignition coil driver electronics, air idle speed control driverelectronics, power down control electronics, and a standardcommunications port. In addition to controlling the engine valvesthrough the hydraulic actuation system, the engine computer alsotypically sequences engine ignition, fuel injection and OBD (onboarddiagnostics).

[0041] The engine computer preferably utilizes a high performancedigital signal processor (DSP), so that control of all aspects of theengines performance can be attained. The DSP interfaces with all of theperipheral sensors, and calculates fuel parameters, ignition timing andengine valve timing based upon prior mapping of the engine. Mapping isperformed multi-dimensionally using engine speed, manifold pressure,induction mass flow and temperatures. In this manner the engine can becontrolled so as to provide maximum fuel economy, minimum emissions,maximum engine torque, or a compromise between these parameters.

[0042] An alternate mapping method to simplify system complexity andreduce parts count would be induction mass flow, temperatures,barometric pressure, engine speed and pedal position sensors.

[0043] The engine computer will determine if the current operatingconditions are within or not within the normal driving cycle of theengine, and will adjust the operation as is required. Configurationsoftware is utilized that allows the reciprocating valve actuation andcontrol system to be tailored for an exact engine system. Engines can bemapped on any engine dynamometer, and evaluated across engine speed andload, so that independent maps can be developed for fuel economy,emissions or torque. Maps are stored for ignition, fuel control andvalve control and can be used separately or in combination.

[0044] The crankshaft position sensor is used to provide the enginecontrol unit with a method of controlling engine valve/fuelinjection/ignition events. The engine crank position sensor must bereliable, accurate, low cost and have a long life. The accuracy andrepeatability should ideally be better than or equal to that of aconventional mechanical camshaft, and with a simple electrical interfaceto the engine control unit. Analog and digital rotational positionsensors can meet these requirements.

[0045] Most analog position sensors can be eliminated if they have anycontacting parts that wear out. Resolvers and sin/cosine (hall effect)potentiometers have output signals that must be phase decoded,digitized, and then require a table lookup to generate a digital angleoutput. These analog sensors usually suffer from long term drift orlinearity/drive problems. A digital sensor eliminates these problems,and is available at low cost. Two types of position encoders are in wideuse today: magnetic (hall effect), and optic (photoelectric).

[0046] Both of these position encoder types are generally available asabsolute position encoders. In addition, an automotive sensor shouldalso be inexpensive and readily mounted to an engine crankshaft. Atypical engine crankshaft has up to ±0.003 inch of axial end play, butgood axial rotational concentricity. Absolute position encoders need tohave precision end play and axial alignment and need to be mounted in avibration and shock free environment to give accurate readouts.

[0047] A 360 count, sin/cosine optical encoder can meet all of the aboverequirements, because recent optical encoder array sensor developmentsallow the encoder to be mounted on the crankshaft and function well inan automotive environment. A magnetic encoder can also be used, but thispresently requires a larger space, and presents somewhat greaterdifficulty to initially index the sensor on the crankshaft for propersynchronization of the engine in an automotive environment.

[0048] For either magnetic or optic encoders, the sin/cosine & indexpulses must be converted into a shaft angle output to control valves,fuel injection, and ignition. It is also desirable for the positionsensor to be able to operate in 2, 3, 4, 5, 6, 8, 10, 12 or 16 cylinderengines; therefore the sensor output counts must be divisible by 2, 3,or 5 to give the same timing to all cylinders (without odd offsets whichcause vibration and uneven operation). This requirement eliminates a 256or 512 count/rev encoder and their simple base 2 arithmetic. With a 360count encoder, a resolution of ¼degree and accuracy of about ⅓degree isobtained from the quadrature output decoding of the sin/cosine signals(and the count is divisible by 2, 3, or 5).

[0049] The engine computer must make valve timing/fuel injection andignition timing computations (or lookup tables) that ensure enginehorsepower/RPM/torque requirements and clean combustion for the engine.Since the engine computer is busy checking many other sensors thatensure clean combustion and efficient operation, it is desirable to“unload” the engine computer by controlling valve timing, fuelinjection, and ignition timing with fixed hardware circuits. Thisunloading also will allow a smaller and lower cost microprocessor to beused in the engine control unit.

[0050] It is desirable to allow the engine computer to give valve timingand ignition or fuel injection updates to the valve control circuits atany time during the engine rotation without risk of damage to valve orpiston position. This becomes more apparent in 8 to 12 cylinder engines,since more events occur during the same engine revolution and atdifferent times than in 4 or 6 cylinder engines. An update to any engineparameter is effective during the current and subsequent control eventsuntil the next update occurs. Thus, the engine computer will not delayupdates until a “safe” point in the cycle is reached to update timingevents. Especially if a cylinder misfires, it is necessary to changesomething immediately if gross pollution is to be avoided, and theengine computer may shut that cylinder off if necessary.

[0051] Engine starting and stopping are a problem using a sin/cosineencoder. During start (power application), the engine sensor does notdetermine its absolute position until the first index pulse is received.Further, at engine shutoff, power will be removed that prevents furthervalve control, so all valves must be quickly closed (for furtheruncontrolled engine rotations). These shutdowns can be easily handled bythe sensor and/or the engine control unit. During a controlled shutdown(ignition switch turned off), valves and engine ignition can be fullycontrolled until zero rotation by the engine computer, sequentiallyshutting off fuel, then closing intake valves, then closing exhaustvalves, then turning off power to itself and engine position sensor.This can be handled with minimum pollution, if desired, or any otherrequirement.

[0052] In case of other, sudden, unexpected power failures, the enginecomputer will shut valves (uncontrolled) with a power fault detectcircuit and local power hold up capacitor. This will prevent enginedamage, and contain most pollutants within the engine.

[0053] During power application (and engine cranking), the engineposition sensor immediately loads default starting values for allvalve/ignition/fuel injection settings. When engine cranking begins, theengine position sensor will command all valves to close (in case any areopen). The engine position sensor will not command and output eventsuntil the first sine/cosine index pulse is received (so absolute crankposition is known). The vehicle driver may have to crank the engine upto one full revolution before this occurs (with all valves closed), butthis will assure adequate hydraulic pressure for a good clean start. Theengine computer may update default engine starting values at any timeafter power application.

[0054] The engine position sensor must also be able to handle reverseengine rotation (safely) if the engine accidently rotates backwards, (ifparked on a hill or during a misfire at startup). These conditions occuronly occasionally, but in all cases, valves must be closed when thepiston is at or near top dead center (TDC) to prevent engine damage.This is performed as a result of standard quadrature decoding.

[0055] The valve actuation and fuel control system software is a fullyinterrupt driven control system that is centered around a DSP processoras a real time engine controller. The valve actuation and interruptsystem software is written in the DSP processor's native instruction setfor speed and efficiency. The other engine sensors operate independentlyfrom the processor, and their routines can be written in a higherlanguage such as BASIC or C⁺⁺, for example.

[0056] The valve actuation and fuel control system can operate bothsynchronously as well as asynchronously with respect to engine rotationintervals. The major operating tasks such as data acquisition anddigital filtration are performed asynchronously in constant timeintervals, but the calculation of some engine parameters, particularlyfuel injection and valve angles, are calculated during degree basedintervals.

[0057] The valve actuation and fuel control system contains a real-timemonitor that allows another software package to query the valveactuation and control system for “while running” information. Thisfeature allows dynamic data updates to be done by another host computersystem for emissions, diagnostic and custom tuning work.

[0058] The valve actuation and fuel control system interfaces to theengine position decoder via an 8 or 16 bit word. This interface setsindividual registers within the decoder, that define starting andstopping points for events in degrees. The degree based eventscontrolled by the valve actuation and engine control system is ignitiondwell, engine valve open position and engine valve closed position ofall intake and exhaust valves as well as the start of the fuel injectionevent. In addition, the start of the fuel injection event is timed suchthat the end of injection event will occur approximately simultaneouswith the spark instant. Because the engine ignition is degree based, thedegrees that the ignition coil are held powered is its dwell, and can beheld either at a constant dwell or at a constant coil energy. The latteris the most desirable for lower power consumption and cooler ignitioncoil operation.

[0059] The propagation delay of the engine valves must be taken intoaccount for top performance. This can be accomplished as part ofvalve/ignition/fuel injection mapping, but as the system ages, and somevalve velocity may be lost, the valve actuation and control system canmeasure its own average valve velocity and recommend a tuneup.

[0060] The valve actuation and fuel control system controls the fuel bysetting the individual injector time periods proportional to the amountof fuel calculated by the engine computer. The start of each injectorpulse can be set at any crank angle and can run for times up to 720crank degrees. The valve actuation and fuel control system can run theinjectors in true sequential or phased sequential patterns for betteratomization. This system could also operate a direct injected gasolineengine.

[0061] With reference to FIGS. 2-7C, the hydraulic valve actuators ofthe reciprocating valve actuation and control system are preferablyprovided as self-contained cartridge assemblies. The hydraulic actuatorspreferably include an actuator piston 60 coupled to the poppet valve,and reciprocating between a first, open position and a second, closedposition, in response to flow of the pressurized hydraulic fluid to thehydraulic actuator. The actuator pistons are preferably sized toefficiently move the engine valves against their return spring forces.This sizing is typically determined by a computer design program thattakes into account all of the necessary mechanical and hydraulicvariables. An ideal piston size is generally one that distributes halfof the pressure drop to the electro-hydraulic valve, and the other halfof the pressure drop to the piston area for actuation. As will beexplained further below, the actuator strokes are preferably terminatedwith hydraulic dampers to assure soft seating of the engine valves.

[0062] As is illustrated in FIG. 2, in one preferred embodiment of thehydraulic actuator of the reciprocating valve actuation and controlsystem of the invention, the actuator piston 60 is mounted to the engine62 by bolts 64. The hydraulic actuator assemblies include a main sleeve66 and a secondary sleeve 68, and the actuator piston is disposed withinthe bore 70 of the main sleeve and the bore 72 of the secondary sleeve.Each of the main and secondary sleeves have precision lapped bores thatmate with the outside diameter 74 of the actuating piston. In addition,each sleeve contains secondary bores 76 that fit closely with a damperland 78 of the actuator piston. The bores and the piston diameters areall concentric, typically with very close tolerances on the order ofplus or minus 0.00005 inch (0.00125 mm). The hydraulic actuator pistonpreferably includes a hydraulic damper system for limiting the actuatorpiston stroke to assure soft seating of the actuator piston, and toavoid overshoot during the engine valve opening stroke and the returnstroke. The secondary bore 76 of the main sleeve therefore defines adamping cavity 80, and the actuator piston includes a damping orifice 82to decelerate the moving parts to avoid overshoot during the enginevalve opening stroke. The secondary bore also preferably defines adamping cavity 84, and the actuator piston includes a damping orifice 86to decelerate the system to avoid high impact of the engine valve intothe valve seat on the return stroke. The stepped land 78 enters thesesecondary diameters in the damping cavities at the ends of the openingand closing strokes, and the oil trapped in the respective cavitiesexits through the respective orifices, thus creating a controlled highback pressure, slowing down the motion of the piston and bringing themoving parts of the valve to a soft landing. Conventional engine valvereturn springs are used as a return device, so that energy stored in thespring drives the closing stroke, and so that energy for the closingstroke does not need to be supplied by the pumping system.

[0063] As is illustrated in FIGS. 3 and 4, in a second embodiment, theactuator piston 90 is mounted in the engine 92 within an alignment tube94, sealed within the engine by the o-ring 95. The actuator pistoncartridge assembly includes a main sleeve 96 disposed within thealignment tube and having a bore 100 mated to the outside diameter 104of the actuator piston. The secondary sleeve of the piston assembly ofFIG. 2 is replaced in this embodiment by the damping ring 106 disposedwithin the alignment tube, and a damping spacer 108. The damping spaceris preferably drilled to provide a gap 110, and is disposed within thealignment tube between the main sleeve and the damping ring. Theactuator piston assembly is preferably contained either as a shrink fitor a pressed fit in the alignment tube. The inside diameter of the mainsleeve can easily be formed to be matched to the outer diameter of theactuating piston, while the outside diameter of the actuating piston canbe sized while on a mandrel that is concentric to the inner bore of thesleeve. These considerations allow the manufacturing cost of theactuator piston and the main sleeve to be relatively inexpensive.Similarly, the damping ring 106 is preferably configured as a bushing,and can easily be manufactured to close tolerances and perfectconcentricity. The damping spacer is also preferably manufactured as abushing, and the gap provided by 110 provides limits for the undampedportion of the stroke of the actuating piston. The orifices 120 providethe damping. The inside diameter of the damping spacer must fit closelyto the damping land 112 on the actuator piston, and the outside diameteris preferably concentric and sized as an interference fit with thealignment tube. However, concentricity and sizing for these closetolerance fits are easily obtained at low manufacturing costs withmodern machining. The alignment tube is preferably manufactured fromprecision tubing, and is preferably made from a seamless tube that iseither honed or roller swaged to size to fit the surrounding bushingparts. The main sleeve, the damping spacer, the damping rings and theactuating piston are preferably preassembled, and are preferably eitherpress fit or shrink fit into the alignment tube. Once in place andchecked for free action, the ends of the alignment tube are typicallyroller swaged or electron beam spot welded to permanently lock the partsin place. The resulting assembly can then be handled as a cartridge, andmounted in the engine with a sealing plug 115, o-ring 114, and a snapring 116. A damping cavity 118 is provided between the outside diameterof the actuator piston and the inside diameter of the damping spacer108, and damping orifices 120 are provided on either side of the dampingland 112 of the actuator piston.

[0064] Referring to FIGS. 5A, 5B, and 6, in another embodiment, theactuator piston 90′ has been modified to replace the stepped actuatingpiston land shown in FIG. 3, in order to reduce manufacturing costs ofthe actuating piston, by allowing the actuator piston to be manufacturedas a cylindrical ground or lapped part. The actuator piston 90′ ismounted in the engine 92′ within an alignment tube 94′, sealed withinthe engine by the o-ring 95′. The actuator piston cartridge assemblyincludes a main sleeve 96′ disposed within the alignment tube and havinga bore 100′ mated to the outside diameter 104′ of the actuator piston.The damping ring 106′ is disposed within the alignment tube, and adamping spacer 108′ that is preferably drilled to provide a gap 110′ isdisposed within the alignment tube between the main sleeve and thedamping ring. The actuator piston assembly is preferably containedeither as a shrink fit or a pressed fit in the alignment tube. Theinside diameter of the damping spacer must fit closely to the dampingland 112′ on the actuator piston, and the outside diameter is preferablyconcentric and sized as an interference fit with the alignment tube. Thealignment tube is preferably manufactured from precision tubing, and ispreferably made from a seamless tube that is either honed or rollerswaged to size to fit the surrounding bushing parts. The main sleeve,the damping spacer, the damping rings and the actuating piston arepreferably preassembled, and are preferably either press fit or shrinkfit into the alignment tube. Once in place and checked for free action,the ends of the alignment tube are typically roller swaged or electronbeam spot welded to permanently lock the parts in place. The resultingassembly can then be handled as a cartridge, and mounted in the enginewith a sealing plug 115′, o-ring 114′, and a snap ring 116′. A dampingcavity 118′ is provided between the outside diameter of the actuatorpiston and the inside diameter of the damping spacer 108′, and a dampingorifice 120′ is provided through the side of the damping land 122′ ofthe actuator piston.

[0065] As is shown in FIGS. 5A and 6, the stepped land of the actuatorpiston can be replaced by a hardened split ring 122′, and the actuatingpiston can be machined with a groove to accept this ring. Since theoutside diameter of the actuating piston is a straight cylinder, theactuator piston can be centerless ground, roller lapped, or otherwisemachined as a straight rod. The hardened split ring is a low cost partthat has a closely sized outside diameter to fit closely to the dampingspacer 108′. The inside diameter of the ring is not critical, and can befit with a high clearance to the actuating piston groove. The hardenedring is typically machined, notched, heat treated, finished to size, andthen is slipped onto a tapered mandrel and split at the notches. The twoparts are kept as a pair and assembled to the actuating piston duringassembly with the alignment tube. One or more damping orifices 120′,such as a multiplicity of holes, slots, flats, and the like, arepreferably formed in the ring, although only a single orifice is shownin FIG. 5B.

[0066] As is illustrated in FIGS. 7A, 7B, and 7C, in another embodiment,the actuator piston 90″ is assembled in the actuator piston cartridgeassembly with an alternative type of replacement of the damping land ofthe actuator piston of FIGS. 2 and 3. The actuator piston 90″ is mountedin the engine 92″ within an alignment tube 94″, sealed within the engineby the o-ring 95″. The actuator piston cartridge assembly includes amain sleeve 96″ disposed within the alignment tube and having a bore100″ mated to the outside diameter 104″ of the actuator piston. Thedamping ring 106″ is disposed within the alignment tube, and a dampingspacer 108″ that is preferably drilled to provide an orifice 110″ isdisposed within the alignment tube between the main sleeve and thedamping ring. The actuator piston assembly is preferably containedeither as a shrink fit or a press fit in the alignment tube. The insidediameter of the damping spacer must fit closely to the damping land 112″on the actuator piston, and the outside diameter is preferablyconcentric and sized as an interference fit with the alignment tube. Thealignment tube is preferably manufactured from precision tubing, and ispreferably made from a seamless tube that is either honed or rollerswaged to size to fit the surrounding bushing parts. The main sleeve,the damping spacer, the damping rings and the actuating piston arepreferably preassembled, and are preferably either press fit or shrinkfit into the alignment tube. Once in place and checked for free action,the ends of the alignment tube are typically roller swaged or electronbeam spot welded to permanently lock the parts in place. The resultingassembly can then be handled as a cartridge, and mounted in the enginewith a sealing plug 115″, o-ring 114″, and a snap ring 116″. A dampingcavity 118′ is provided between the outside diameter of the actuatorpiston and the inside diameter of the damping spacer 108″, and dampingorifices 120″ are provided on either side of the damping land 112″ ofthe actuator piston.

[0067] In this embodiment, the actuator piston damping land is replacedby a sealing ring, such as a two turn laminar sealing ring, such as aSmalley laminar sealing ring. Such a ring is generally available frommanufacturers of spiral snap rings at a relatively low cost. Either one,two or three of these rings typically can be assembled into theactuating piston groove. The radial spring action of the ring keeps therings in contact with the damping spacer 108″, thus assuring lowhydraulic fluid leakage. Small holes can also be drilled through theserings to act as one or more damping orifices 120″, one of which is shownin FIG. 7B. Alternatively, the damping orifices in the actuator pistonof FIG. 2 can be used. An advantage of using the laminar sealing ringsis that the bore in the damping spacer can have a much relaxedtolerance, and all that is necessary is that a reasonably smooth surfacebe provided.

[0068] With reference to FIGS. 8-15, the electrically operatedelectro-hydraulic valves are generally of a rotary design. Theelectro-hydraulic valves 28 provide multiple paths for flow of thehydraulic fluid, such that the sum of the open areas in the valve islarge, and relatively small rotational angles switch the cylinder portsfrom a pressure supply configuration to a return path configuration.Referring to FIGS. 8-11, the electrically operated electro-hydraulicvalves preferably include a rotor or rotary valve element 130, assembledin combination with a three path latched magnetic motor 132.

[0069] The rotor is provided with a pressure supply groove 134 thatcommunicates with a plurality of axial pressure grooves 136 that branchfrom the pressure supply groove 134 and dead-end. A second set of axialreturn grooves 138 is also provided in the rotor, communicating at theopposing end of the rotor with the return to the system via the engineoil sump, and are dead-ended at their ends adjacent to the pressuresupply groove. The rotor is preferably manufactured of high strength,hardened steel or an equivalent durable material. The outside diameterof the rotor is typically machined to a high finish and is precisionsized to fit within the stator, or fixed valve element 140.

[0070] With reference to FIGS. 8 and 12-15, the stator is preferablyprovided with an inlet pressure port and an inner bore 144, with whichthe inlet pressure port is in fluid communication through a plurality ofradially oriented holes 146. The stator also includes a cylinder portgroove 148 in fluid communication with the inner bore and the axialgrooves 136 and 138 of the rotor through a plurality of axial statorslots 150. The stator is also preferably fabricated of high strength,hardened steel or an equivalent durable material, and the insidediameter is also typically machined to a high finish and precision sizedto mate with the rotor. The stator is installed in a housing 152 thatprovides the necessary fluid connections with the pressure supply andpressure return lines of the hydraulic fluid system, and the rotaryvalve housing 152 is assembled together with the housing 154 of themagnetic motor assembly.

[0071]FIGS. 14 and 15 show the rotor and stator mated for operation,with FIG. 15 illustrating how the pressure will be distributed, in thevalve cross-section. As can be readily appreciated, alternate grooves ofthe rotor will be either pressurized with the supply of pressurizedhydraulic fluid, or will be at return pressure, depending upon theorientation of the rotor within the stator. The cylinder ports 150 arevented to the return grooves 138, and when the rotor is turned,preferably 9% clockwise, the cylinder ports will be connected to thepressure grooves 136. A hydraulic actuator connected to the cylinderport will then receive flow from six pressure grooves.

[0072] The open flow area of the valve depends upon the axial length ofthe cylinder port slots, and the diameter of the rotor-stator interface.The electrically driven magnetic motor assembly, connected to the rotor,can thus on command rotate the rotor first clockwise, and thencounterclockwise, 9%. Other angles of rotation may, of course, also besuitable. It should thus be apparent that the rotary valve can open avery high flow area when rotated through relatively small angles. Ifadditional area is required, the rotor and stator can be designed withincreased length and the stator provided with longer cylinder portslots, as desired. In this manner, the valve design can be adapted to avariety of applications. The rotor design also inherently provides avery small rotational mass moment of inertia, since the numerous grooveson the outside diameter of the rotor have removed a substantial amountof material mass that would otherwise contribute to rotational inertiaof the rotor. The small rotational angle required for operation of therotary valve, and the low mass moment of inertia of the rotary valveboth optimize the operation of the reciprocating valve actuation andcontrol system of the invention for operation at very high cyclic rates,with a low power consumption by the electrical actuator.

[0073] The rotation of the cylindrical rotor element also entails verylow friction, since the radial loading on the rotor is pressure balancedat all times, so that wear on the rotor and stator of the rotary valvewill be minimized. It should be readily appreciated that the rotaryvalve design could easily be modified to provide a return passagesimilar to that used for the inlet pressure port, and an elongatedversion could also include a secondary group of cylinder ports to createa four way valve. It should also be readily appreciated that the rotorand stator are ideally configured for manufacture by investment castingor metal injection molding methods, which will permit greater economy inthe manufacturing process.

[0074] Referring to FIGS. 8, 9, 16 and 17, the electrically operatedelectro-hydraulic valves preferably are provided with a rotating motordriver capable of fast response to electrical pulses, with magneticlatching at two positions. Briefly, the magnetic motor consists of athree path magnetic circuit, with each of the three paths meeting at acentral point. Two of the magnetic paths pass through individualmagnetizing coils, while the third path includes a rotor and astationary permanent magnet that holds or latches the rotary element inthe position last commanded by the engine computer.

[0075] As is best seen in FIG. 16, the first path of the magnetic motoris comprised of a first pole piece 160, connected to a firstelectromagnetic coil 162 energized by the electrical signals from theengine computer, and the magnetic junction 164 connected to the firstpole piece and first coil. The second path of the magnetic motorsimilarly is comprised of a second pole piece 166 connected to a secondelectromagnetic coil 168 energized by electrical signals from the enginecomputer, and the magnetic junction 164 to which the second pole pieceand second electromagnetic coil are connected. The third path of themagnetic motor is comprised of the magnetic rotor 170 mounted forrotation between a first position and a second position contacting thefirst pole piece and second pole piece, respectively, an air gap 172between the magnetic rotor and a third pole piece 174, a permanentmagnet 176 connected to the third pole piece, and a fourth pole piece178 connected between the permanent magnet and the magnetic junction164. A rotary output shaft 180 is provided on the rotor of the magneticmotor for transferring the rotary motion of the rotor of the magneticmotor to the rotor of the rotary valve 130. Referring to FIG. 17, thefirst and second pole pieces are preferably arranged to form 30% gaps atthe end of the rotor of the magnetic motor, to provide maximum leverageand maximum torque. When the rotor is attracted to either the first polepiece or the second pole piece, the gap between the rotor and one of thepole pieces closes, creating a minimum reluctance path, and thepermanent magnetic flux in the third path of the magnetic motor latchesthe rotor of the magnetic motor in place, as indicated by referencenumber 182.

[0076] The operation of the magnetic motor will be further describedwith reference to FIGS. 16 and 17. If the permanent magnet is orientedto produce a north pole at the rotor of the magnetic motor, at rest,both the first and second pole pieces would be at south polarity. Thelatched position then completes the permanent magnet flux path, suchthat the north polarity end of the rotor is magnetically latched to thesouth polarity of the pole piece which the magnetic rotor contacts. InFIG. 17, the magnetic rotor is shown latched to the first pole piece160, so that in order to move the magnetic rotor from the position shownto latch with the second pole piece, the second electromagnetic coil 168is pulsed with direct current. The current flow in the secondelectromagnetic coil is preferably phased to produce a strong south poleat the second pole piece 166. When this occurs the second pole pieceattracts the magnetic rotor, and since the first pole piece is on theopposite end of the magnetic path of the second electromagnetic coil168, the first pole piece assumes a north polarity. Since the magneticrotor is permanently provided with a north polarity by the permanentmagnet, the magnetic rotor is repelled from the first pole piece, and isattracted to the second pole piece. At the same time, the north polarityflux from the second electromagnetic coil 168 enters the third paththrough the junction 164, reinforcing the permanent magnet, andstrengthening the north polarity of the rotor. The magnetic rotor isthen very strongly urged to close the gap with the second pole piece,and once this gap is closed, and the coil electrical pulse has ended,the permanent magnetic flux from the third magnetic path latches themagnetic rotor in contact with the second pole piece. If theelectromagnetic coil 162 is then pulsed, the opposite action occurs,with the first pole piece acquiring a strong south polarity, and thesecond pole piece acquiring a north polarity, and the permanent magnetand magnet rotor receiving reinforcement of the north polarity. Thesecond pole piece then repels, and the first pole piece attracts themagnetic rotor, and the permanent magnet again latches the magneticrotor to the new position at the first pole piece. As should be readilyapparent, the permanent magnet may also be installed to produce a southpolarity at the rotor, at which both of the electromagnetic coilsrequire current flow phased to produce north polarity at the first andsecond pole pieces. The resulting functions will then be the same asdescribed above, with all of the magnetic polarities described reversed.

[0077] Testing of the three path rotary latched magnetic motor has shownthat the motor is capable of very high speed operation. With a rotationcyclic angle of 9%, cyclic rates of 260 Hertz can be achieved with 12volts, 5 ampere electrical pulses of 1.0 ms duration (0.06watt-seconds). At the 260 Hertz rate, the magnetic motor drew a steadyoperational current of 1.172 RMS amperes.

[0078] The improvement in the speed of operation of the reciprocatingvalve actuation and control system of the invention can be readilyappreciated with reference to FIG. 18, comparing valve speeds of amechanical camshaft driven engine and the camless engine valve controlsystem of the invention. The graph shows the length of the valve strokein inches vs. degrees of rotation of a mechanical camshaft. Whengraphed, the cycle of opening and closing of a valve driven by amechanical camshaft will display a shape similar to a sine curve. Theperiod (as measured in crankshaft degrees) remains constant for anyengine load or rpm. However, the cycle of opening and closing of valvesdriven by the reciprocating valve actuating and control system of theinvention operates much faster. Designed to match valve opening rates atthe maximum engine rpm, the valve actuation and control system of theinvention opens the valve at this same rate regardless of engineoperating conditions. Thus, the valve actuation and control system ofthe invention will match the valve rate at a maximum rpm of an engine,but will be faster at all lesser engine speeds. Because of this improvedspeed, the reciprocating valve actuation and control system of theinvention allows greater flexibility in programming valve events,allowing for improved low end torque, lower emissions and improved fueleconomy.

[0079] The reciprocating valve actuation and control system has theability to alter the valve cyclical stroke number (i.e., 2 cycle) to adesired valve cycle combination. It is therefore conceivable to startand run an engine in standard 4 cycle mode, then change over at sometime to 2 cycle mode and effectively double the potential availabletorque.

[0080] The reciprocating valve actuation and control system also has theability to control the effective engine speed without the use of athrottle valve. This is accomplished by controlling the valve durationfrom its idle duration to its maximum torque duration as a function ofthe desired throttle position. This allows simplification of theinduction system and allows for a more compact engine design. Thethrottle control abilities also provide the ability to control anengine's volumetric efficiency under certain conditions, and allow theengine to have a softer RPM limiting function.

[0081] Upon sensing ignition switch shutoff of system power failure, thereciprocating valve actuation and control system and valve spring putsthe valve in the most desirable “generally closed” state, so that thevalve positions are not ambiguous and will thus protect engines fromvalve/valve or piston valve contact. After the valve positions areguaranteed, the reciprocating valve actuation and control system willturn off the power to itself, and operations will cease.

[0082] The stored energy in the accumulator can be used for engine powerbursts. During these brief power bursts, the hydraulic pump can bedisengaged, allowing the valves to be powered solely from stored energyfrom the accumulator with additional energy savings derived by notoperating the hydraulic pump. Also, during braking, some energy thatwould normally be absorbed by the vehicle friction braking system can bestored in the accumulator. This is possible because the crankshaft(ultimately) is connected to the vehicle wheels and can drive thehydraulic pump to fill the accumulator for future hydraulic valveactuation.

[0083] A controller chip can eliminate the need for a half crankshaftspeed cam position sensor along with all of its mechanical andelectrical interfaces. (Typically the distributor or cam positionsensor.) The chip can calculate and determine overlap and firingsequencing of a 2, 4, 5, 6, etc cycle engine during the start-upsequencing.

[0084] While the preferred embodiment describes the use of engine oilfrom the engine lubrication circuit, an alternative would be a secondaryfluid (e.g. diesel fuel, ATF, steering fluid, etc.). The hydraulic fluidmay be also be a separate system with another fluid type on a separatefluid circuit. Also, the fluid return reservoir may be the enginecrankcase, or a separate and different location.

[0085] By use of the invention, multiple intake or exhaust valves of acylinder need not open at the same time. A delay of even a small amountcan off-load the driver electronics and reduce peak current load. Thiswill allow smaller current traces on the circuit board and preventringing of the power transistors. The delay of the intake valves openingin a multi inlet valve cylinder can enhance the swirl effect. Bothopening and closing events of the set of valves can be mapped to enhancevarious operating characteristics. This effect can also be combined withthe use of shaped and directed inlet ports. The invention can alsoenhance mechanical simplicity of the intake system. Installing a PedalPosition Sensor at the velocity/accelerator pedal will allowsimplification of the induction system by eliminating throttle platesand effectively throttling the engine using only the conventional intakeand exhaust valves that open into the cylinder.

[0086] Since the invention allows broad control of a variety ofcombination functions, an internal EGR function can be created bycommanding a second set of exhaust valve opening and closing eventsduring the intake sequence. Similarly, the intake valve may be openedand closed several times during the intake or exhaust sequence topromote scavenging and later to follow the piston to promote intakevolumetric optimization, and intake and exhaust valves may be ditheredto control engine throttling and braking.

[0087] As a further indication of the benefits of the invention, oneintake port would be designed for high swirl (lower volume) while asecond intake port would be designed for high volume (lower swirl).During throttled conditions, only the high swirl port would be used tooptimize combustion efficiency. If exhaust valves are provided asdifferent sizes, the smaller would be opened first so as tosubstantially lower cylinder pressure prior to opening the secondexhaust valve. When both valves are of equal size, either valve could beopened ahead of the second to again lower cylinder pressure beforeopening the second valve. This sequencing may allow the use of smallervalve actuators and certainly reduced energy to operate the secondvalve. Engines with both multiple intake and exhaust valves can be madeto operate under higher conditions of swirl. Although paired intake andexhaust valves may be of equal size, swirl is maximized by havingdifferent sized valves and properly sequencing them. Refer to FIG. 19.Sequence is as follows:

[0088] a. #1 Intake valve 184 opens (largest valve)

[0089] b. #2 Intake valve 186 opens (smaller valve)

[0090] c. #2 Intake valve 186 closes

[0091] d. #1 Intake valve 184 closes

[0092] e. Compression and power stroke take place.

[0093] f. #4 Exhaust valve 190 opens (smaller valve w/less surface area)

[0094] g. #3 Exhaust valve 188 opens (larger valve w/more volume)

[0095] h. #3 Exhaust valve 188 closes

[0096] i. #1 Intake valve 184 opens (overlap begins)

[0097] j. #4 Exhaust valve 190 closes (overlap ends)

[0098] The invention can also effectively use a bridge in the combustionchamber to assist swirl. In addition to valve size and sequencing topromote higher swirl, the upper combustion chamber may incorporate a“bridge” effectively separating the intake side from the exhaust side inthe dome of the combustion chamber. With the “bridge” in place, gaseswould be better directed to flow in a “swirl” pattern as shown in FIG.19.

[0099] Using the invention, engines having multiple intake or exhaustvalves could be start sequenced having only one intake and one exhaustvalve operating. The invention permits reprogramming to allow reverseengine rotation by simply inverting one input wire pair. Reverseoperation is advantageous to operation of marine equipment having dualoutdrives or T-drives, since vehicle torsional accelerations arecanceled by reverse rotational engines. This feature would alsoeliminate the need for reverse gear(s) in the transmission since forwardgears would be used to operate in either vehicle direction. Thisprovides an opportunity for multiple reverse gears without addedhardware.

[0100] It will be apparent from the foregoing that while particularforms of the invention have been illustrated and described, variousmodifications can be made without departing from the spirit and scope ofthe invention. Accordingly, it is not intended that the invention belimited, except as by the appended claims.

What is claimed is:
 1. A reciprocating valve actuation and controlsystem for the cylinders of an internal combustion engine, comprising: apoppet valve moveable between a first and second position; a source ofpressurized hydraulic fluid; a hydraulic actuator including an actuatorpiston coupled to the poppet valve and reciprocating between a first andsecond position responsive to flow of the pressurized hydraulic fluid tothe hydraulic actuator; an electrically operated valve controlling flowof the pressurized hydraulic fluid to the actuator; and control meansgenerating electrical pulses to control the electrically operated valve.2. The reciprocating valve actuation and control system of claim 1 ,wherein the electrically operated valve controlling flow of thepressurized hydraulic fluid to the actuator supplies pressurizedhydraulic fluid to the hydraulic actuator when electrically pulsed to afirst position, and dumps pressurized hydraulic fluid to a system returnwhen electrically pulsed to a second position.
 3. The reciprocatingvalve actuation and control system of claim 2 , wherein the electricallyoperated valve comprises a three path rotary latched magnetic motor. 4.The reciprocating valve actuation and control system of claim 1 ,wherein the electrically operated valve comprises: a rotary valve havinga housing; a stator having an inlet pressure port receiving pressurizedhydraulic fluid, an inner bore in fluid communication with the inletpressure port through a plurality of radially oriented apertures; acylinder port groove in fluid communication with the hydraulic actuator;a plurality of axial slots formed in the stator allowing fluidcommunication between the cylinder port groove and the inner bore of thestator; a generally cylindrically shaped rotor disposed within thestator, the rotor having a pressure supply groove at one end forreceiving pressurized hydraulic fluid from the inlet pressure port ofthe stator; a plurality of axial pressure grooves in fluid communicationwith the pressure supply groove of the rotor for supplying pressurizedhydraulic fluid to the actuator; and a plurality of return groove formedin the rotor in fluid communication with a pressurized hydraulic fluidreturn, for receiving hydraulic fluid from the hydraulic actuator. 5.The reciprocating valve actuation and control system of claim 3 ,wherein the three path rotary latched magnetic motor comprises: a firstpole piece connected to a first electromagnetic coil energized byelectrical pulses from said control means; a second pole piece connectedto a second electromagnetic coil energized by electrical pulses fromsaid control means, said first and second pole pieces being connected ata magnetic junction; a magnetic rotor disposed for rotation between afirst position and a second position contacting said first and secondpole pieces, respectively; a third pole piece disposed adjacent to themagnetic rotor so as to define an air gap between the magnetic rotor andthe third pole piece; a permanent magnet connected to third pole piece;a fourth pole piece connected between the permanent magnet and themagnetic junction; and an output shaft mounted on the magnetic rotoroperatively connected to rotary valve means for controlling flow of thepressurized hydraulic fluid to the hydraulic actuator.
 6. Thereciprocating valve actuation and control system of claim 1 , whereinsaid hydraulic actuator comprises a self-contained cartridge assemblyincluding an actuator piston having means for damping a stroke of theactuator piston to assure soft seating of the actuator, and to avoidovershoot of the actuator piston.
 7. The reciprocating valve actuationand control system of claim 6 , wherein said means for damping comprisesfirst damping means to avoid overshoot during an opening stroke of theengine valve.
 8. The reciprocating valve actuation and control system ofclaim 7 , wherein said means for damping comprises second damping meansto decelerate the actuator piston to avoid high impact of the enginevalve into the valve seat.
 9. The reciprocating valve actuation andcontrol system of claim 6 , wherein said means for damping comprises astepped land on the actuator piston.
 10. The reciprocating valveactuation and control system of claim 6 , wherein said self-containedcartridge assembly further comprises a main generally tubular sleevehaving a bore, said bore having a surface defining a damper cavity, saidactuator piston having a damper land member, and said damper cavityreceiving said damper land member during an actuating stroke of saidactuator piston, whereby hydraulic fluid is trapped in the damper cavityto damp motion of the actuator piston during a stroke of the actuatorpiston.
 11. The reciprocating valve actuation and control system ofclaim 10 , further comprising a secondary generally tubular sleevehaving a bore, said secondary sleeve bore having a surface defining asecondary damper cavity, and said actuator piston having a surfacedefining a damper orifice for fluid communication of said hydraulicfluid from one of said main sleeve damping cavity and said secondarysleeve damping cavity to the hydraulic fluid return.
 12. Thereciprocating valve actuation and control system of claim 10 , when saidself-contained cartridge assembly further comprises an alignment tubewithin which said main sleeve is disposed, a generally tubular dampingspacer disposed within said alignment tube adjacent to the main sleeve,a damping ring disposed within said alignment tube adjacent to saiddamping spacer, and said actuating piston having a surface defining adamping orifice for fluid communication of hydraulic fluid from saiddamper cavity to the hydraulic fluid return.
 13. The reciprocating valveactuation and control system of claim 12 , wherein said damper landmember comprises a split ring, said split ring having a surface defininga damper orifice through said split ring for communicating hydraulicfluid to the hydraulic fluid return.
 14. The reciprocating valveactuation and control system of claim 12 , wherein said damper landmember comprises a laminar sealing ring, said sealing ring having asurface defining an orifice in the sealing ring for communication ofhydraulic fluid to the hydraulic fluid return.
 15. The reciprocatingvalve actuation and control system of claim 1 , wherein said source ofpressurized hydraulic fluid comprises an engine driven hydraulicpositive displacement pump for supplying said hydraulic fluid pressure,said hydraulic fluid is engine oil, and an engine oil sump connected influid communication with said pump for supplying engine oil to the pump,and said engine oil sump being connected in fluid communication forreceiving return engine oil from the valve actuation and control system.16. The reciprocating valve actuation and control system of claim 15 ,further comprising an unloader valve connected in fluid communicationwith the pump for limiting output pressure of the pump.
 17. Thereciprocating valve actuation and control system of claim 16 , furthercomprising a check valve to prevent backflow from the accumulator. 18.The reciprocating valve actuation and control system of claim 16 ,further comprising an accumulator connected in fluid communication withthe pump and the unloader valve for storing a volume of the hydraulicfluid.
 19. The reciprocating valve actuation and control system of claim16 , wherein said unloader valve comprises a pressure sensing valve forsensing pump output pressure, said unloader valve opening when the pumpoutput pressure reaches a preset threshold value, said unloader valvereturning flow of said hydraulic fluid to return.
 20. The reciprocatingvalve actuation and control system of claim 1 , wherein said controlmeans comprises a computer and a plurality of sensors disposed in theengine for sensing engine variables, and optimizing performance of thereciprocating valve actuation and control system.
 21. The reciprocatingvalve actuation and control system of claim 1 , the internal combustionengine having a cylinder head and a combustion chamber, and wherein theengine cylinder head has a bridge dividing the combustion chamber. 22.The reciprocating valve actuation and control system of claim 1 ,wherein said control means comprises a digital signal processor to takeadvantage of its high speed real time signal processing capability,whereby crankshaft dynamic related problems are diagnosed, and dealtwith in real time.
 23. A method for controlling reciprocating valveactuation for the cylinders of an internal combustion engine in areciprocating valve actuation and control system, the system including apoppet valve moveable between a first and second position; a source ofpressurized hydraulic fluid; a hydraulic actuator including an actuatorpiston coupled to the poppet valve and reciprocating between a first andsecond position responsive to flow of the pressurized hydraulic fluid tothe hydraulic actuator; an electrically operated valve controlling flowof the pressurized hydraulic fluid to the actuator; and an enginecontrol unit gene rating electrical pulses to control the electricallyoperated valve, wherein the source of pressurized hydraulic fluidcomprises an engine driven hydraulic positive displacement pump forsupplying the hydraulic fluid pressure, an unloader valve connected influid communication with the pump for limiting output pressure of thepump, and an accumulator connected in fluid communication with the pumpand the unloader valve for storing a volume of the hydraulic fluid, themethod comprising the step of: storing hydraulic energy in theaccumulator.
 24. The method of claim 23 , further comprising the stepof: controlling the accumulator in a way that commands the engine drivenpump to “run free” or be disconnected during brief power bursts.
 25. Themethod of claim 23 , further comprising the step of: controlling theaccumulator in a way that forces the accumulator to be charged duringbraking.
 26. The method of claim 23 , further comprising the step of:controlling the accumulator in a way that forces the accumulator to becharged during the time the vehicle needs to decelerate.
 27. A methodfor controlling reciprocating valve actuation for the cylinders of aninternal combustion engine in a reciprocating valve actuation andcontrol system, the system including a poppet valve moveable between afirst and second position; a source of pressurized hydraulic fluid; ahydraulic actuator including an actuator piston coupled to the poppetvalve and reciprocating between a first and second position responsiveto flow of the pressurized hydraulic fluid to the hydraulic actuator;and an electrically operated valve controlling flow of the pressurizedhydraulic fluid to the actuator; the method comprising the step of:controlling the electrically operated valve with an engine control unitgenerating electrical pulses.
 28. The method of claim 27 , wherein saidstep of controlling comprises: controlling the engine control unit in away that commands a delay to take place in the opening of multipleintake or exhaust valves in the cylinder.
 29. The method of claim 27 ,further comprising the step of: the engine control unit controlling thevalve timing to create a swirl effect in the combustion chamber.
 30. Themethod of claim 27 , further comprising the step of: mapping the enginecontrol unit in a manner that optimizes the swirl effect.
 31. The methodof claim 27 , wherein said step of controlling comprises: the enginecontrol unit controlling the valve timing of the intake and exhaustvalves of an engine having at least three valves per cylinder, such thatthe intake and exhaust valves will not open at the same time, andcontrolling the valve timing of the intake and exhaust valves of theengine to provide a delay to off load driver electronics and reduce peakcurrent load, allowing smaller current traces and preventing ringing ofpower transistors.
 32. The method of claim 27 , the engine having amulti-inlet valve cylinder having shaped and directed inlet ports,wherein said step of controlling comprises: the engine control unitcontrolling the valve timing to provide a delay of the opening of intakevalves, to cause a swirl effect to take place that is augmented by theshaped and directed inlet ports.
 33. The method of claim 27 , the enginehaving a multi valve cylinder having first and second exhaust valves,and first and second hydraulic actuators, the second exhaust valve beinglarger than the first exhaust valve, the first exhaust valve to openbeing smaller in head diameter, resulting in lower actuation pressure,wherein said step of controlling comprises: the engine control unitcontrolling the timing of the valves to create a delay between theopening point of exhaust valves in the multi valve cylinder to reducethe demand placed on the second actuator, to lower horsepower requiredto drive the larger exhaust second valve.
 34. The method of claim 27 ,the engine having four intake and exhaust valves, wherein said step ofcontrolling comprises: the engine control unit controlling the timing ofthe valves in the following sequence: a. number 1 Intake valve opens(large valve) b. number 4 Exhaust valve closes (after start up) c.number 2 Intake valve opens (smaller valve) d. number 2 Intake valvecloses e. number 1 Intake valve closes f. compression and power stroketake place g. number 4 Exhaust valve opens (smaller valve w/less surfacearea) h. number 3 Exhaust valve opens (larger valve w/more volume) i.number 3 Exhaust valve closes j. number 1 Intake valve opens (overlapbegins) k. number 4 Exhaust valve closes (overlap ends).
 35. The methodof claim 27 , the engine control unit commanding a first set of exhaustvalve opening and closing events, wherein said step of controllingcomprises: the engine control unit controlling the timing of the valvesby commanding a second set of exhaust valve opening and closing eventsto take place.
 36. The method of claim 27 , the engine having fourintake and exhaust valves, wherein said step of controlling comprises:the engine control unit controlling the timing of the valves in thefollowing sequence: a. number 1 Intake valve opens (largest valve) b.number 2 Intake valve opens (smaller valve) c. number 2 Intake valvecloses d. number 1 Intake valve closes e. compression and power stroketake place f. number 4 Exhaust valve opens (smaller valve w/less surfacearea) g. number 3 Exhaust valve opens (larger valve w/more volume) h.number 3 Exhaust valve closes i. number 1 Intake valve opens (overlapbegins) j. number 4 Exhaust valve closes (overlap ends).
 37. The methodof claim 27 , wherein said step of controlling comprises: the enginecontrol unit controlling the valve timing by opening and closing thevalves several times during the same stroke.
 38. The method of claim 27, wherein said step of controlling comprises: the engine control unitcontrolling the valve timing by opening and closing the valves severaltimes to control throttling and braking.
 39. The method of claim 27 ,wherein said step of controlling comprises: the engine control unitcontrolling the valve timing.
 40. A method for controlling reciprocatingvalve actuation for the cylinders of an internal combustion engine in areciprocating valve actuation and control system, the system including apoppet valve moveable between a first and second position; a source ofpressurized hydraulic fluid; a hydraulic actuator including an actuatorpiston coupled to the poppet valve and reciprocating between a first andsecond position responsive to flow of the pressurized hydraulic fluid tothe hydraulic actuator; an electrically operated valve controlling flowof the pressurized hydraulic fluid to the actuator; a crankshaft; and acontrol means controlling the electrically operated valve, the methodcomprising the step of: determining the position and direction ofrotation of the crank shaft from the electrical outputs of a sin/cosinecrankshaft position sensor.
 41. The method of claim 40 , the methodfurther comprising the step of: operating valve openings and closingsthat are correct for forward/reverse crankshaft rotation, based upon thecrankshaft position and direction information, to eliminating possiblemechanical interference for crankshaft reverse rotation.
 42. The methodof claim 40 , further comprising the step of: reversing the directionindication by electronically inverting one signal to cause the engine torun backwards.
 43. The method of claim 40 , the engine being a fourcycle engine having a distributor and camshaft, the method furthercomprising the step of: dividing the electrical position output of theelectrical crankshaft position sensor by two electronically to eliminatecostly mechanical components that drive the distributor and camshaft athalf speed, and determining the initial timing during startup sequencingfor valves, fuel and ignition.
 44. The method of claim 40 , the methodfurther comprising the step of: inputting preset default valve, fuel andignition operating values into registers of the control means uponapplication of power, to be utilized if the control means fails tooperate, allowing operation in emergencies.
 45. The method of claim 44 ,further comprising the step of: the control means closing any openvalves upon application of power to eliminate mechanical interferenceuntil correct crankshaft location is determined by a startup sequence.46. The method of claim 40 , further comprising the step of: the controlmeans inhibiting fuel injection but not ignition during a shut-offcommand by the operator.
 47. The method of claim 46 , further comprisingthe step of: sequentially commanding all intake and exhaust valves ofthe engine to close before power termination to the control system, toeliminate any possible mechanical interference after power removal, inorder to provide smooth termination, a low pollution termination, or arapid deceleration termination, depending on the actual valve closureand ignition sequencing.
 48. The method of claim 40 , further comprisingthe step of: comparing whether the electrical crankshaft position anddirection outputs of the position sensor are greater than but not equalto desired values that open valves, to allow for commanded valve eventvalue changes asynchronously to crankshaft position without missedevents, such as a missed valve open event causing a misfire and greatervibrations, noise and emitted pollution problem.
 49. The method of claim40 , further comprising the step of: comparing whether the electricalcrankshaft position and direction outputs of the position sensor aregreater than but not equal to desired values that open valves, to allowfor commanded valve event value changes asynchronously to crankshaftposition without missed events, such as a missed valve closure eventcausing a mechanical interference problem.
 50. The method of claim 40 ,further comprising the step of: comparing whether the electricalcrankshaft position and direction outputs of the position sensor aregreater than but not equal to desired values that open valves, to allowfor commanded valve event value changes asynchronously to crankshaftposition without missed events, such as a missed fuel injection eventcausing a misfire and greater mechanical vibrations and noise.
 51. Themethod of claim 40 , further comprising the step of: comparing whetherthe electrical crankshaft position and direction outputs of the positionsensor are greater than but not equal to desired values that openvalves, to allow for commanded valve event value changes asynchronouslyto crankshaft position without missed events, such as a missed ignitionevent causing a misfire and greater emitted pollution.
 52. A method forcontrolling reciprocating valve actuation for the cylinders of aninternal combustion engine in a reciprocating valve actuation andcontrol system, the system including a poppet valve moveable between afirst and second position; a source of pressurized hydraulic fluid; ahydraulic actuator including an actuator piston coupled to the poppetvalve and reciprocating between a first and second position responsiveto flow of the pressurized hydraulic fluid to the hydraulic actuator; anelectrically operated valve controlling flow of the pressurizedhydraulic fluid to the actuator; and a control means controlling theelectrically operated valve, the method comprising the step of:controlling the dynamic performance of the system by newly addeddimensions of mapping strategy, consisting of existing mapping ofsensory inputs such as gas pedal position, inlet and exhaust manifoldand barometric pressures, exhaust gas composition, coolant and ambientair temperatures with ignition timing and fuel injection, along with newdimensions of inlet/exhaust valve timing.
 53. The method of claim 52 ,the engine having individual fuel injectors, the method furthercomprising the steps of: controlling the valve actuation solenoids; andcontrolling the individual fuel injectors, as well as individual sparkevents on each cylinder used on the system, based upon said sensoryinput, and based upon multidimensional mapping.
 54. The method of claim52 , further comprising the step of: determining if any cylinders areoperating unsatisfactorily, based upon use rotational rate measurements.55. The method of claim 54 , further comprising the step of: the controlmeans disabling defective cylinders entirely, reducing pollution andpotential further engine damage, while offering a limited “limp home”operation.